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  • From time to time, a heat ex-changer is designed carefully yet fails to achieve the desired performance by a wide mar-gin, achieving, say, only half

    the duty. With an understanding of some of the more common reasons why this might happen, design-ers can avoid these problems in the first place, and troubleshooters can recognize the root causes quickly.

    Exchangers for single-phase opera-tion, condensing and boiling are con-sidered in that order here; but as we shall see, exchangers often handle a combination of these, and it is not always obvious which process is caus-ing the problem. In fact, some of these problems are quite unexpected and can even take experienced designers by surprise.

    It must be recognized that the most important cause of problems in ex-changers is excessive fouling. Other articles, books, and conferences have been dedicated to this problem, so fouling will not be addressed here. Instead we consider those exchangers that have failed for some reason other than fouling.

    SINGLE PHASEGood flow patterns are keyIn a single-phase exchanger, most problems arise when the unit is de-

    signed to meet an unrealistically high thermal effectiveness. See Realistic Expectations, p. 45, for the upper lim-its of common heat exchangers. Oc-casionally, people inadvertently try to exceed these values. While not neces-sarily impossible, such an approach really pushes ones luck.

    To achieve realistically high ther-mal effectiveness, countercurrent flow is normally essential. Also, both streams must be distributed evenly across any flow path and among all parallel flow paths. Furthermore, there must be no axial mixing. Flow that achieves these characteristics to-gether is normally referred to as plug flow. Plug flow, while necessary, is not a sufficient condition on its own to en-sure high thermal effectiveness. All parallel flow paths must also undergo identical heat transfer processes.

    In general, shell-and-tube heat ex-changers are not good at achieving plug flow and identical heat transfer in parallel paths, which is why they

    cannot achieve as high a thermal ef-fectiveness as other types. The main problems arise on the shell side.

    In trying to achieve 90% effective-ness with a shell-and-tube exchanger, it is essential to ensure good arrange-ment of nozzles and headers so that equal flow occurs in all tubes. On the shell side, care is needed in the choice of baffle pitch and cut to minimize re-circulation paths behind baffles and thereby minimize the resultant axial mixing. It is very important to avoid bypass flows. It may even be necessary to go to a special design like a twisted-tube exchanger. For a practical example of this scenario, see Example 1, p. 45.

    The Z and U flow arrangementsPlate-heat-exchanger manufacturers know well that of the two flow ar-rangements shown in Figure 1, the U arrangement gives a better flow distribution in the plate pack than the Z arrangement. One might think that the Z arrangement would be bet-

    Feature Report

    Sometimes fouling is not the problem. To get closer

    to your design duty, consider these practical

    design tips

    Feature Report Feature Report Feature Report

    44 CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004

    Heat Exchanger Duty: Going for Gold

    David ButterworthHeat Transfer and Fluid Flow Service

    FIGURE 1. In the plate exchanger with a Z arrangement, momentum changes throughout the exchanger give rise to a larger (unwanted) variation between its header pressures (graphs above and below each diagram) than that of a U ar-rangement. Therefore, the U arrangement is preferred for its more identical behav-ior of parallel channels

  • ter, since it gives equal-length flow paths for each parallel stream. (See Example 2, p. 47, for an illustration of how misleading this concept can be.) However, we have to look at the mo-mentum changes in the header as well as the frictional losses.

    The small graphs over and under each figure show the pressure change in the headers when momentum

    change effects are taken into account. It can be seen that the Z arrangement gives rise to a larger variation in the pressure difference that drives the flow in the parallel channels.

    CONDENSATIONVentingOne of the main causes of problems with condensers is the failure to vent

    noncondensable gases. This results in the depression of the dewpoint in the condenser and hence the loss of tem-perature driving force. High noncon-densable-gas concentrations also lower the heat transfer coefficient. Consider these rules-of-thumb for good venting: Vent from the cold end of the con-

    denser, where noncondensable con-centrations are highest

    Avoid any direct paths (from inlet to vent) that do not cross any tubes

    Keep the pressure drop per unit length as uniform as possible along the flow path to drive the noncon-densable gas toward the vent (thus avoiding formation of noncondens-able gas pockets)

    Keep the vent clear of the conden-

    EXAMPLE 1. SHELL-AND-TUBE GAS HEATER

    CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004 45

    This case involves a cross-flow shell-and-tube exchanger with two tube-side passes formed with U tubes (TEMA AXU type). The shell side has nitrogen gas, which is to be heated from 21.5C to 135C by condensing steam at 151C inside the tubes. Under the design operating conditions, almost all the thermal resis-tance is on the shell side around 98%. While this arrangement is apparently a simple design, a quick calculation shows that the thermal effectiveness is 88%, a high enough value that any prob-lems arising will have serious effects.

    The original design was as shown in Figure A (right). However, during construction of the plant, problems were found with the pip-ing layout so that the nozzles and bundle layout were rearranged as shown in Figure B. Nobody thought to block the interpass by-pass flow path on the shell side, so the exchanger performed far below the design duty. The outlet gas temperature achieved was only 97C as compared with the design value of 135C. The actual heat load was only 66% of the design value, and the overall heat transfer coefficient appeared to be only 42% of the design value.

    The first attempt at a solution was to immediately insert a metal box to block the interpass gap, as shown in Figure C. Unfortu-nately, while some improvement was achieved, the problem was not completely solved. The gas outlet temperature was raised to 116C from 97C but still fell short of the design value of 135C. The overall heat-transfer coefficient had increased to 66%.

    At one point, the designers decided to replacing the square tube layout with a triangular configuration to increase the number of tubes. It is good that this option was not actually implemented, because it would not have solved the problem.

    The problem was instead with the condensing on the tube side. The way the U tubes were arranged meant that there was really a single pass on the tube side rather than two passes. To achieve the design condition, the first tubes encountered by the gas flow would experience a temperature difference of 129.5C, and the last only

    16C. This is a ratio of 8.1 to 1. As the design overall coefficient was constant, the same ratio was being expected for the heat loads to the first and last tubes.

    In the final and ultimately successful solution, the tube bundle was rotated as shown in Figure D. Calculations proved that to achieve the design condition, the outer U tubes would have to handle about twice the duty of the inner U tubes. This seemed better than before, but we were not sure if this alone would solve the problem. So we went for both a rotation of the bundle and a change to a triangular pitch. The solution was successful, giving an outlet gas temperature of 140C and thermal effectiveness of 91.5%, an improvement for both values over the design value.

    Incidentally, after all of the attempts to troubleshoot this ex-changer, it would seem that the original design (Figure A), would have worked perfectly.

    5"#-&46((&45&%."9*.6.7"-6&40'5)&3."-&''&$5*7&/&44

    Exchanger type .BYJNVN Shell-and-tube 90%

    Plate-and-frame 95%

    Plate-fin 98%

    Printed circuit 98%

    REALISTIC EXPECTATIONS

    One important cause of problems is when the exchanger has been designed to have an unrealistically high thermal effectiveness (). The definition of thermal effectiveness is

    (1)

    where Q is the heat added or removed from the stream, and Qmax is the theoreti-cal maximum amount of heat that can be added or removed.

    The most important stream in this context is the one with the highest value of . For single-phase streams with constant specific heat, Equation (1) becomes, for the hot and cold stream, respectively (of a two-stream exchanger):

    (2a)

    (2b)

    When one is trying to achieve a high ther-mal effectiveness, any flaw in exchanger performance will have serious conse-quences. Some exchangers are capable of higher thermal effectiveness than others. Table 1 indicates the maximum percentage that can be realistically achieved for differ-ent types. To achieve the values in Table 1, great care is required in design. Going be-yond these values is not recommended.

  • sate layer to prevent flooding Be sure that parallel flow paths

    have identical dutiesFigure 2 shows a vertical, shell-side condenser with a sensibly positioned vent that is at the cold end of the condenser and high enough up not to flood with condensate. It has been known for people to put the vent at the top of the shell, using some mistaken reasoning that noncondensable gases rise. The problem with that arrange-ment is that the whole of the con-denser must fill with noncondensable gas before the vent starts to operate. More information on the good loca-tion of vents is given elsewhere [1]. Here are explanations of some slightly unexpected venting errors that have caused serious problems.

    Parallel condensing paths with different dutiesFigure 3 illustrates a problem that can occur with air-cooled condensers. In this particular case, the condenser has one tube-side pass and two rows, but the problem can occur with other ar-rangements as well. In the illustrated case, the temperature difference of the bottom row is greater than that of the top row. Because of the difference in heat loads, all of the vapor entering the bottom row condenses, while some in the top row does not. In compensa-tion, the bottom row sucks in the vapor from the outlet header that has not been condensed in the top row. As the

    vapor enters the bottom row from both ends, the noncondensable gases cannot be vented and will be trapped, causing a deterioration in performance of the bottom pass.

    The problem of parallel paths with different duties can occur in many situations (also illustrated in Exam-ple 1). Additionally, Figure 4 shows a TEMA J-type shell with condensation on the shell side. The problem in this arrangement is that there is only one tube-side pass for the coolant flow. Thus, one half of the exchanger has a higher heat load than the other, caus-ing a pocket of noncondensable gas to form away from the vent. Incidentally, the vent is actually in a sensible posi-tion for designs with many tube-side passes, which would cause the two halves to have nearly the same duty.

    Zero pressure drop condenser with dead zonesIt is possible, particularly in vacuum condensers, to have zero pressure drop, because the frictional losses bal-ance the momentum recovery that results from vapor deceleration. The problem with this condition, of course, is that there is nothing to drive out the gas. Increasing the venting is not a good solution, because rather than clearing the dead tube, it tends to suck out vapor from the operating tube. The problem may not manifest itself at first, since time is needed for the noncondensable gases to accumulate. In fact, the problem could go unno-ticed for years causing a significant loss of thermal efficiency throughout the plant if, say, only one of two con-densers in parallel is affected.

    Feature Report

    46 CHEMICAL ENGINEERING WWW.CHE.COM DECEMBER 2004

    FIGURE 2. The placement of vents in vertical, shell-side condensers should be at the cold end of the condenser, just high enough to avoid condensate flooding

    FIGURE 3. A lesser heat load in the top row of a parallel path inhibits complete condensation of vapor, which ultimately limits performance of the bottom pass

    FIGURE 4. With only a single tube pass for coolant flow, this exchanger has a higher heat load on one half than on the other. This causes uneven condensation and inefficient heat transfer

    FIGURE 5. With this configuration, the large vapor inlet flow and rapid condensa-tion achieve a large enough deceleration and pressure recovery to overcome the frictional pressure drop. To be stable, the zero-flow tube has to be filled with non-condensable, inert gas

  • So, with vacuum operation, never design a condenser to have zero pres-sure drop. Also, follow the good vent-ing rule that the pressure drop per unit length through the condenser is kept as uniform as possible. For con-densers in parallel, always have sepa-rate vents, rather than relying on a manifold system.

    In Figure 5, for instance, all flow is running through the bottom tube. With this configuration, the large vapor inlet flow and rapid condensa-tion achieve a large enough decelera-tion and pressure recovery to over-come the frictional pressure drop. To be stable, the zero-flow tube has to be filled with noncondensable, inert gas.

    Condensate not drainingCondensation gives a high heat-trans-fer coefficient, whereas a stagnant pool of condensate gives a very low heat-transfer coefficient for the tubes sit-ting in the pool. So, if the condensate is not draining properly, the efficiency of heat transfer will be compromised. The problem is somewhat common during plant startup, because debris that is generated during construction gets lodged in the outlet line. If that is not the cause, the problem may be arising because the condensate outlet line is too small.

    Failure to drain condensate is a problem that sometime occurs with vertical thermosiphon reboilers heated with condensing service steam. Usu-ally, the reboiler has been designed with an unrealistically high fouling resistance, so that, initially, while clean, it will over-perform.

    To bring the reboiler back to the de-sired performance, the common reac-tion is to reduce the steam pressure, which in turn lowers the condensing temperature. However, the steam trap

    may not have been designed to oper-ate at the lower pressure. If not, the condensate will not drain properly, and the shell will fill with condensate until the pressure from the liquid head is high enough to force the steam trap to work. This problem manifests itself periodically, as a slow deterioration in performance and a jump back to good performance. Such unstable behavior in the reboiler can be violent enough to cause instabilities in the distillation column to which it is attached.

    BOILINGUnpredictable nucleate-boiling heat...